Rotary hydraulic coupling of the turbine type



Nov. 4, 1947. E. A. THOMPSON 2,430,258

ROTARY HYDRAULIC COUPLING OF THE TURBINE TYPE Original Filed Feb. 5, 1940 3 Sbeets-Sheet 1 w j Immnio'u 647] Q. Thain 175022 73g 2 W gflomeys Nov. 4, 1947. E. A. THOMPSON 2,430,258

ROTARY HYDRAULIC COUPLING OF THE TURBINE TYPE Original Filed Feb. 5, 1940 3 Sheets-Sheet 2 I 500 I000 IVG/NE E. I? M 'Jmwnto v 54 11 a, Thom 250m i zww atto'msus Nov. 4, 1947. E. A. THOMPSON ROTARY HYDRAULIC COUPLING OF THE TURBINE TYPE Original Filed Feb. 5, 1940 3 Sheets-Sheet 3 attorneys Patented Nov. 4, 1947 ROTARY HYDRAULIC COUPLING OF THE TURBINE TYPE Earl A. Thompson, Bloomfield Hills, Mich., as-

signor to General Motors Corporation, Detroit, Mich., a corporation of Delaware Original application February 5, 1940, Serial No.

317,348; now Patent No. 2,357,295, dated September 5, 1944. Divided and this application February 17, 1941, Serial No. 379,248

8 Claims. (01. 60-54)- The present invention relates to power devices for transmitting variable torque through turbine bladed rotors placed adjacent each other, and. operating by toroidal circulation of fluid in a closed toroidal turbine circuit, and more particularly in power devices for motor vehicles. This is a divisional application for Letters Patent of my United States application Serial No. 317,- 348, filed February 5, 1940, now Patent No. 2,357,295, for improvements in Turbine drive mechanisms.

A feature of the present invention is the conflning of the fluid power transfer media in a toroidal vaned space, constituting a substantially closed flow circuit for the fluid, without voids or compression and decompression Zones.

Another feature of the present invention is the provision of a vane turbine rotor structure in which the ratio of the number of vanes to the volumetric space occupied by the vanes them-- selves is high, this ratio factor being a measure of fineness, from which important gains in emciency and low stalling torque are derived. The vanes may be of given flexibility for mounting the mass of the ring structure of the supporting rotor members.

A further feature is the provision of a method of assembly of such fluid turbines utilizing slotted sheet metal stampings of semi-toroidal shape in the slots of which tabbed vanes are inserted, the structure being fastened together by rolling the tabs of the vanes aforesaid, with even pressure applied circumferentially, with mechanical force, supplemented by welding where needed.

The vanes may be of flexible, stamped sheet metal for the reason above stated.

An additional feature is the provision of fluid circuit-closing axial spacer members at the hub portions of the assembly of the rotors of the invention, which clamp the vanes axially into position.

A supplementary feature is the provision of a constantly replenished body of fluid moving dynamically through the turbine device, supplied with means for compelling circulation from and to a supply reservoir having a large heat radia tion capacity, while the circulation system is equipped with means to furnish and maintain a given pressure of operating fluid in the working spaces of the turbine device, thereby elimi:

nating voids, frothing and gasification of the fluid.

The terminology used further herein to describe the turbine members or elements will name that one of the members coupled to the source 2 of power as the impeller and that one coupled to the load shaft as the rotor."

These objects and features of the invention, in combination, and additional features further described and combined with them, contribute to the attainment of the objects afore-mentioned, and other objects as appear herein.

An individual embodiment of my invention, which is defined in its full aspects in the appended claims, is shown in the accompanying drawings.

Figure 1 is a longitudinal elevation section of the fluid turbine of the invention with parts broken away, showing the fluid circulation system.

Figure 2 is a perspective view of a sector cut from one of the turbine elements.

Figure 3 is an exploded view of the method of assembly of the vanes of the turbine arrangement, as they appear in Figure 1.

The chart of Figure 4 is a typical power curve of the device of the invention.

Figure 5 is a vertical section of a transmission driving assembly embodying the invention as described in Figure 1. Figure 5a is a schematic outline of the transmission assembly of Figure 5 showing the transmission brake actuators.

The engine shaft 1 in Figure 1 is connected to flywheel 2 to which is bolted drum 3 attached to hollow shaft 4 supported in bearing 6 of housing l0. Shaft 4 has aflixed gear I and may drive clutch and gearing mechanism at the right, not involved directly in this invention. The inner solid shaft l3 may be driven directly by said engine-connected drum 3, through gearing at varying speed ratios thereby, or be coupled to shaft l as required for a particular drive and as will be understood from the description of- Figure 5.

The inner end of hollow shaft 8 is splined at 9 to accommodate hub member M, the flange of which supports impeller 20 and spacer ring l2; Solid shaft I3 is the output member of the driving assembly, and may be connected directly to output or through gearing cooperating with the gearing aforesaid.

Rotor turbine hub I4 is splined to shaft l3, its flange mounting rotor elements 2| and spacer ring I6.

Spring loaded check valve 22 in passage 23 drilled in hub l4 permits excess fluid pressure within the operating spaces of the turbine device to drain to passages 24, 25, and 21 in shaft l3, thence through the various bearings along shaft I3 back to the lubrication sump. Spring 22a 3- may be set for a predetermined pressure relief value.

Vertical shaft 28 supported in extensions of easing l carries gear 29 meshing with gear I of hollow-shaft 4 and drives pump assembly 30 located in the lower portion of the casing so as to draw lubricant from the sump. The pump is driven, therefore, directly by the engine.

Pump feed passage 3| in housing ||la delivers oil through the passages 32, 33, and 34 to space 35 so that the turbine impeller and rotor elements and 2| are surrounded with oil and kept filled and under pressure at all times, the check valve 22 relieving excess pressure as described. It will be understood that while engine-connected shaft 4 is rotating, there will be a constantly changing supply of oil for lubrication and for furnishing the operating liquid of the turbine elements 20 and 2|. The supply thus recirculating tends to reach a mean operating temperature level quickly and remain there during the running of the mechanism, for the whole sump and casing ID of the transmission driven by shafts 4 and I3 are adapted to act as radiators of excess heat. It will be noted that the pressure within drum 3 provided by the pump is exerted on the outer flanks of the rotors l1 and Ila to counteract that of the working space which tends to separate them axially.

While it is commonly known to utilize the zonal differences in operating pressures within fluid work turbines for maintaining and relieving the pressures, such devices as injectors and aspirators are subject to variations in the turbine operating conditions, and it is preferred therefore to provide'po'sitive circulation of fluid held under pressure established by the operation of the engine and connected parts. My method in practice has been found to avoid the common troubles with voids, overheating, ga-sification and frothing of the liquid, and the consequent interference with proper operation and life of the mechanism involved.

Attention is directed to the fact that spacer rings I2 and I6 close off the low velocity spaces of the turbine elements, difiering from the constructions commonly used, and providing a substantially closed path for toroidal circulation of fluid within the turbine elements.

It should be observed that the pressure values within the fluid turbine working space will be subject to constant variation because of the fact that in a vehicle, for example, the power re-- quirements set up a need for speed variation of the engine. This, in turn, sets up a variation in the pump pressure made available, so that the fluid in the working space is circulating practically all the time, whenever the check valve 22 is being opened and closed.

This effect is, of course, accentuated by leakage through the various bearings of the fluid rotor hubs and shafting.

Turbine impeller element 20 is composed of external semi-torus shell I! and internal member l8, slotted radially to receive vanes I9 displaced as shown in Figure 2. Similarly, rotor element 2| is made up of shell Ha, internal member la and vanes |9. The elements 20 and 2| are identical.

Each vane I9 is cut as shown in Figure 3 with tabs 4|, 42, 43, and 44 projecting so as to pass through the slots 36-38 of the shells and slots 31 of the inner members as indicated. The vanes may be cut from sheet metal of predetermined flexibility to cooperate with the mass of the members I8, |8a for torsional vibration dampening purposes.

Each of the external shells I1 and In has an external circumferential bead 40 better shown in Figure 3, behind which the tabs 4| of the vanes I9 are bent after being inserted in slots 36 and 38. The tabs 4| are finally rolled flat to the outside contour of the shells 2|l2| and may be brazed, soldered or Welded, as strength and rigidity require, although rolling will sufllce for all practical purposes.

The inner member l8 of the assembly is half ring-shaped, resembling half of a torus, its section, however, being of special contour, for preserving efliciency in guiding the fluid flow around the eye of the turbine member.

The piece |8 is slotted radially at 31 from the closed side of the ring to a depth equal to the lateral spacing between the registering inner and outer edges of the vanes so that tabs 42 and 43 may be bent and rolled smoothly within the reentrant portion of the piece. This slotting is shown in the right-hand portion of Figure 3.

' The tabs are rolled circumferentially in one hand of rotation to preserve dynamic balance. The inner circumferential grooves 46 and 41 accommodate the bent tabs 42 and 43, which after rolling. or other affirming operation, present a relatively smooth surface.

At the inner portion of the shells 2|l2|, the tabs 44 are mounted in radial slots 38 cut in hubs II and M, the tabs being locked against endwise movement by spacer rings i2 and i6. Bolts 5| thread into the rings Hi and |2, holding the shells Ho and I! to the hubs I4 and H. Bearing 52 supports hub l2 on shaft |3.

The method-.of construction. provides exceptional lightness and strength when compared with competitive devices. It has been found in practice that it is notnecessary to weld the parts together if sufficient care is taken to dimension the slots and vane thickness, and to use a sheet material for the vanes capable of bein bent and rolled for locking in position as described. This obviates the risk of heat warping the accurately aligned parts. It has been found in practice that the stresses in the members are negligible so that after assembly, there is no internal warp of the impeller and rotor elements 20 and 2|, which assures eflicient operation for an indefinite period. The trueness of the dynamic balance of the turbine elements is assured by this method, so that noises caused by precession of the rotating masses or other variations in the smooth flow of torque are not experienced.

In. practice, dynamic unbalance of a rotor assembly may be corrected by weighted tabs such as shown at 50 at Figure 2, spot welded to inner ring IE or I 8a, or to outer shell IT, or Ila, such as indicated at 5m in Figure 3. It is desirable that these units be free from dynamic unbalance. The yield of the flexible vanes may be correlated to the mass values of the rotating parts such as the semi-toroidal rings l8, |8a.

It should be understood that the presence of check valve 22 in hub |4 removes material which disturbs the perfect balance of the rotor assembly Ila, which unbalance may be rectified by utilization of weights such as shown in Figure 2 at 50.

The highly accurate construction afforded by the present invention enables the designer and user to make extensive use of the present fluid turbine device in that with the present construction, extremely high speeds of engine and load 5. shafting may be safely handled. The vanes are all cut from absolutely uniform sheetstock from a common die, and are therefore of equal weight for dynamic balance.

Attention is directed to the important fact that the toroidal space enclosed between shells H and Ila and inner members I8, I81: is completely vaned within that space, and further, that the circumferential boundary of the space is not interrupted or broken at any point, except at the median plane bisecting the torus. This continuity of flow characteristic is extremely important if losses from turbulence and interruption of unitary cross-section flow in the vane pockets is to be avoided. Figures 1 and 2 show clearly the complete sectional filling of the space between shells l1 and I8 by the vanes l9. The hub portion of Figure 1 is equipped with spacer rings I2 and I6 so that the circumferential flow is not altered. This feature is herein stressed in View of the many disclosures in the prior art which show the vanes of similar devices cut away at their inner terminals with the rotor hubs, and which show circumferential voids in the closed fluid circuit about the eye of the torus adjacent the hub.

It has been found by experiment, and by actual experience in a motor car drive that the presence of such voids and interruptions in the liquid flow circuit causes eddies and stagnant areas, as well as compression and decompression effects, all of which reduce the efliciency of the fluid flywheel device. Since energy which is not transmitted from input shaft to output shaft can only be expressed as a heat loss, the present invention, because of the elimination of the above noted undesirable features, makes an important contribution to the art. It makes the closed-circuit fluid turbine clutch for the first time a practical reality. The efliciency as indicated by the torque curve, such as indicated in Figure 4, rises very sharply, regardless of the rate of acceleration of the engine, with no appreciable lag, so that at a speed of 1000 R. P. M. input, for example, the

device of the example is capable of transmitting the full torque of the average automobile engine.

This means that the car driver can have the engine throttle wide open, with the fluid flywheel transmitting full power at high efficiency, at any practical engine speed above 1000 R. P. M., for example.

It will be noted that the supply system provided by pump 39 first fills the fluid turbine working space to rated pressure before the check valve 22 furnishes lubrication pressure to the transmission bearings beyond passages 32, 33, 34 and space 35. This eliminates air pockets which may have formed while the device may have been at rest.

Figure is a vertical elevation section of a complete transmission drive unit embodying the present invention. The general transmission arrangement is that of United States application Serial Number 267,024, filed April 10, 1939, now matured as Letters Patent No. 2,211,233, issued August 13, 1940, to O. K. Kelley.

The construction at the left of the figure is the same as in Figure 1.

Central shaft I 3 extends to the right, terminating in a pilot bearing 49 in the socket of output shaft 50, and having attached or integral sun gear 5|. Surrounding hollow shaft 8 has attached to its carrier flange 52 for planet gears 53, and extends to the right, terminating in splined clutch hub 69.

The hollow shaft 4, transmitting engine power, is integral with annulus gear 54.

The gearing shown is arranged in three groups, a front unit, a rear unit, and a reverse unit. The front .unit is composed of meshing annulus 54, planets 53, and sun gear 55, the latter being splined to drum 56 of brake 60. The spindles 51 for planets 53, extend through from carrier flange 52 to clutch hub 58, splined to receive clutch plates 6|. The interior circumferential portion of drum 56 is splined to carry mating clutch plates 62.

The rear unit is composed of meshing sun gear 5|, planets 63,, and annulus gear 64, the drum 65 of which is splined internally for clutch plates 61, mating with clutch plates 68 mounted on hub 69, attached to hollow shaft 8. The external surface of annulus drum 65 is braked by servo-actuated brake 10. The carrier 1 I for planets 63 is affixed to, or integral with the transmission output shaft 50.

The reverse unit is composed of meshing annulus gear 13, planets I4, and sun gear I5 fixed to rotate with annulus gear 64 of the rear unit. The carrier 16 for planets I4 is splined to the output shaft 50. Annulus gear 13 is integral with drum 11 supported in bearing 18, and toothed at 19 to engage reverse locking pawl 80, sliding in recess 8| in casing l0. The pawl is released by spring 82 and loaded for engagement with teeth 19 by fluid pressure delivered through pipe I05.

The engine power is applied to the whole assembly through annulus gear 54, constantly rotating with the engine. When brake 60 is applied, sun gear 55 is stopped, furnishing reaction for reduced forward drive of carrier 52, sleeve 8 and impeller 20 of the fluid turbine unit. When the brake 60 is released and the clutch plates 6|,

62 are pressed together, a locking couple is established, causing the annulus 54, sun gear 55 and carrier 52 to rotate as one, providing drive at engine speed to sleeve 8 and rotor 20 of the fluid turbine unit.

Alternate energisat-ion of brake 60 or clutch 6|, 62 provides therefore low speed drive to the turbine unit or drive at engine speed.

The rear unit is capable of drive when either brake 10 or clutch 61, 68 are energised. Assuming that the engine power may be transmitted through the fluid turbine unit to shaft l3 and sun gear 5|, the application of brake 10 will stop annulus gear 64 affording reaction for low speed forward drive of carrier II, and output shaft 50. Release of brake l0, and energisation of clutch 61, 68 causes annulus gear 64 and shaft 8 to rotate as one so that the net rotation of carrier H and output shaft 50 will be the resultant of two components, one derived from rotation of annulus 64 and the other from rotation of sun gear 5|.

The reverse, low speed drive is obtained by energisation of brake 60 of the front unit, and pawl 80, which locks annulus 13, to establish reverse ,gear reaction.

Fluid pressure pipes HH and I02 of gland I00 lead respectively to cylinders 85 and in drums 56 and 65, for applying controlled fluid pressure to pistons 86 and 9| respectively, in order to energise clutch 6|, 62 or clutch 6T, 68. Brake plunger cylinders shown in Figure 5a at 81 and 92 and fed by pipes I03 and I04, house pressure actuated pistons for actuating the brakes 60 and 10 in a commonly known manner. The external control system for setting up a combination pattern of driving speed ratios is not a part of the present invention. The following table of speed ratios is 7 given, however, to provide a clear-cut description of the operation obtainable with the combinations of the present invention; the symbol :1: indicating energisation or actuation of the element so marked:

Front Unit Roar Unit Clutch Brake Clutch Brake The direction of movement of lubrication oil from the check valve porting 24 and 21 of Figure 1 through the shaft passages leading to the transmission elements can be followed in Figure 5. Radial passage 21 of Fig. 1 opens into a space between shaft 13 and shaft 8, from whence flow may be directed from radial passage I06 to the front unit gearing of Fig. 5, from radial passage ID! to the members adjacent gland I00, through passage I08 and along passage N of shaft I3 to passage IH feeding the gearing of the rear unit; and through the pilot bearing 49. The system assures that all of the rotating parts are plentifully supplied with lubricant, which because of the rotation of the gears and drums, is thrown out by centrifugal force to the inner walls of the casing in, whence it drains into the sump to be recirculated by pump 30. The spreading out of the body of oil over the interior of the casing facilitates cooling of the oil, which effect is likewise augmented by the sump pan Nib acting as a radiator. r

The method of assembly of the vane, shell and hub elements provides a light weight, compact unit, which is capable of high speeds such as 3600 to 4500 R. P. M., without dynamic unbalance.

The slots 35 and 38 in the members I! and II are cut accurately to accommodate the thickness of the vane elements with just enough clearance to permit the vanes [9 to be slipped into place. The rolling down of the tabs 41, 42 and 43 anchors the vanes in place and supports the whole structure,which has comparative strength to that of a many-spoked wheel.

In the present invention the number of vane pockets and vanes with respect to the collective thicknesses of the vanes determines a factor which is dcscribable as the fineness ratio. For example, the added thicknesses of all of the vanes when compared with the total of 360 degrees, may amount to 30 degrees. The vanes then take up one-twelfth of the total effective face area of the turbine wheel. If, for example, the number of vanes is 48, the ratio of 48 to 30 is 1.6, a value taken as approaching useful efficiency. Now if there were only 30 vanes, the ratio would be 1.0; and if there were only 24. it would be 0.8, a value believed lacking in useful character. Should there be 60 vanes, the fineness ratio would be 2.0, a preferred ratio over either of the above given.

Now if the additive vane thicknesses consumed only degrees (or one-eighteenth), yet there were 48 vanes, the fineness ratio would be 2.4; and if there were vanes, it would be 1.5; 24 vanes would yield the ratio 1.2, and 60 vanes, a fineness ratio of 3.0. The present invention, because of the method of assembly,- allows the designer to use very thin blades, and as large a number of vanes as the torque capacity characteristic of the design of the required fluid flywheel calls for. The disclosure herein then permits of a relatively high fineness ratio, which yields a minimum of drag loss at the parting plane between the fluid flywheel rotors. This provides a low unit pressure in each vane pocket also.

The use of a relatively large number of thin, flexible vanes allows the designer to use water, or very light oil as the transfer fluid in the fluid flywheel, and provides a design of extremely low stalling torque.

The feature of high fineness ratio in combination with the feature noted above, of the closed circuit toroidal section, produces a result of unusual utility in that heat losses, intolerable in earlier devices in this art, are now reduced to a minimum, while maintaining the desired low stalling torque.

These considerations apply more particularly to usages of fluid turbine couplings wherein efficiency over a wide speed range is desired as distinct from usages in which one eiflcienrw is made deliberately low over a starting speed range in which large inertias are required to be overcome.

In operation, the hollow shaft 8 receives direct rotation from the engine when clutch GI, 62 of the front unit is engaged and indirect rotation from engine connected drum 3 and shaft 4 hen brake 60 is applied and clutch 6|, 62 released, and in either case the fluid in the vane pockets of element 20 begins to circulate radially when the engine is accelerated. Since there is a relatively closed fluid circuit, because of spacers i6 and I2, the fluid tends to move about the eye iii-18a of the toroidal space, as seen in Figure 1. The inner curved face of shell 2!) guides the fluid to first move longitudinally, and the fluid velocity energy is imparted to the element 2|, the fluid returning toward the shaft center as it loses its velocity. At first the resultant rotation of output element and shaft l3 connected to the load is small, and there is a high differential of speed between shafts 8 and i3. As shaft l3 increases in speed, the force generated in the moving torus of fluid increases, therefore the torque capacity of the device as a clutch increases, Figure 4 giving a curve representing an example of the power results obtained.

At some predetermined speed of shaft 8 for a given torque applied to shaft l3, the differential of speed has diminished to a low value, so that the eiflciency of the device has attained a correlated high value. Measured efliciencies show results close to per cent, for given constructions, power plants and loads, in normal motor car driving speed ranges.

Due to the accurate contour of the vane pockets and parts l'l-l8, the fluid passing through from the low to the high velocity space is not subject to voids, or to increases or decreases in cross sectional area of the compartments, therefore there is a minimum of loss ordinarily encountered in such devices resulting from turbulence and skin friction factors.

When the power shaft I is idling it will be seen that pump 30 will likewise idle, only delivering sufiicient fluid under low pressure to keep the working space between the turbine rotors filled, for cflicient operation during the ensuing driving interval. The check valve 22 assists in this action by blocking the outflow of fluid from the working space, by the setting of the check valve spring 22a, so that the pump only has to make up minor leakage losses during the idling period.

It should be made clear that the references in this specification to the resiliency of the rotor construction and its ability to absorb torque impulses or vibrations is due to the fact that with the rolled down vane tabs in the shell and in the core ring, radial tensional stresses in the vanes caused by such impulses, originating in an acceleration or deceleration component applied to one shaft or another, tend to be absorbed if they are of a frequency and energy value within the range of the natural period of the core ring taken as a mass and the vanes as deflecting springs.

It is useful to provide a constant circulatory path for the fluid arranged such that the crosssectional areas at any point in the path are equivalent or equal. Without the closed, continuous low velocity area, provided by the spacer rings lZ-IG, and the described means for shaping the path of flow between the circular external shells 20 2I and their inner members l8l8a, the noted benefits in efliciency derived are not obtainable. l

The further improvement of positive fluid circulation under controlled pressure for both lubrication and cooling, effective at all times when the engine is running, is believed to be of unusual utility. In this connection, it is conceded that ordinary pump feed to fluid turbine devices is old in the art, but the present system of the invention provides two features not heretofore described, namely, the method of sustaining a given pressure level at all times in the turbine, and the exhausting of the spent pressure through lubrication passages ,to a reservoir of extended heat radiation capacity.

The preceding demonstration has disclosed a unique method of confining the working fluid in the turbine in a completely closed and voidless toroidal space; it has shown a structure yielding a high fineness ratio; and a method of assembly from stamped sheet parts which yields excep-- tional lightness, strength and inherent dynamic balance while utilizing the closure members for the toroidal circuit as clamps for the vanes. The structure disclosed lends itself to application of torsional vibration absorption technique as described herein.

The advantages of these disclosures are believed obvious to those skilled in the art, and are thought to represent substantial contributions thereto. It should be understood that the illustrative form is not in any sense restrictive, and that the invention may be employed in many ways within the scope of the appended claims,

Having herein fully described the invention, what I claim as new and desire to secure by Letters Patent is:

1. In power devices for motor vehicles, a power shaft and a, concentric load shaft, fac ng turbine rotor members secured to each of said shafts for transmitting variable torque therebetween, fixed radial vanes in said members each of which extends to the parting zone therebetween, a, toroidal fluid circulation space included between the members about which the fluid may circulate rotationally and circumferentially, a power driven drum enclosing said members providing an input fluid pressure space therefor, hubs for each of said members, spacer elements on each of said hubs extending to said parting zone, for closing the contour of said toroidal space, a fluid pressure supply means connected to the interior of said drum for creating uniform pressure therein tending to oppose axial separation of said members, and control means responding only to the pressure within said space and located in one of said hubs for relieving pressure therefrom at above a predetermined value for preventing axial separation of said members and maintaining the said space completely filled during rotation of said drum.

2. In a fluid coupling for driving and driven members, a flywheel structure secured to said driving member, a drum attached to said flywheel structure adapted to be connected to and disconnected from a concentric hollow shaft, a shaft within said hollow shaft, axially spaced facing fluid rotors attached to each of said shafts, said rotors enclosing a substantially closed toroidal fluid circuit space, hubs for each of said rotors for attachment to said shafts respectively, the inner surfaces of said hubs providing flow guidance in said space, a fluid supply space enclosed within said drum and communicating with the axial space between said rotors, the pressure of said supply space being operative to oppose axial separation of said rotors, fluid pressure supply means driven by rotation of said drum adapted to furnish fluid under pressure to said rotors by delivery to the space enclosed by said drum, and control means responsive to the pressure of said toroidal space for maintaining the working pressure at all times when said supply means is operating.

3. In fluid coupling turbines, coacting axially spaced driving and driven fluid rotor members each rotating with a shaft for transmitting variable torque therebetween, hubs for each of said members for attachment to its respective shaft, a pump driven by means rotating with the driving member, passages connecting said pump with the working space between said members, an external rotatable drum forming a fluid inlet delivery space for the fluid pressure developed by said pump and enclosing said members, said delivery space substantially enclosing the members so as to provide pressure opposing their separation and opening to the axial space between the members, outlet passages in one of said hubs for relief of pressure from the working space between said members, and a check valve located in said hub immediately adjacent the outlet passages and subject only to the pressure developed within the working space of said members for relieving all pressure in said space above a given value, whereby the working space remains fllled at all times when rotation is applied to said pump.

4. In fluid turbine driving devices, a power shaft, a load shaft, a primary shaft, turbine hubs fastened to each of said power and load shafts, coacting fluid rotor members mounted on said hubs for transmitting variable torque between the said shafts and axially spaced to form a parting zone between the members, said members being formed to include a vaned working space, a pump driven by rotation of said primary shaft, a drum surrounding said members and rotatable by said primary shaft at varying speed ratios with respect to the speeds of either of said shafts, the space between the drum and members constituting the pressure delivery space open to said parting zone, pressure delivery passages including passages in said hubs connecting said pump with said drum, and control means mounted in one of said hubs and responding to the pressure of said work ing space for maintaining the fluid working space between said member under a given positive fluid pressure at all times when said pump is rotating, and for opposing the effect of separation of said members by pressure Within said working space.

5. In fluid turbine drive devices, fo-r'motor vehicles, a primary shaftyand a secondary shaft,

such that during operation the said working space generates a high pressure peripheral zone and a, lower pressure central zone the pressure of which tends to separate the members axially, a fluid reservoir, 9, fluid supply means adapted to draw fluid from said reservoir and deliver it under pressure to the working space between said vaned members through a fluid delivery space external to said members, and control means located adjacent th low pressure zone of said working space effective to regulate the degree of pressure of said working space and that tending to separate the members, said fluid control means constantly changing the fluid of the said working space at all times while said primary shaft delivers torque to said secondary shaft.

6. In fluid turbine drive devices, in combination, a power shaft, a load shaft, rotor hubs secured to each of said shafts, coacting fluid rotors mounted to rotate with each hub to clutch the shafts together by turbine action derived from power delivered by said power shaft, a vaned fluid turbine pressure working space between said rotors in which axial force tending to separate them is generated, a check valve mounted in one of said hubs and connected to said working space, a circuit of fluid pressure supply for maintaining the working space between said rotors under super-atmospheric pressure comprising a pump driven by rotation of said power shaft, suction passages to said pump, pressure passages connecting the pump with said working space, pressure relief passages leading to said check valve and from said check valve to lubrication passages including passages in one of said hubs, a reservoir serving as a supply for said pump, and means regulating the operation of said check valve effective to establish a constantly active circulation of fluid through said circuit when said pump is operating, and to limit the said axial pressur tending to separate said rotors.

7. In power transmission devices, driving and driven shafts, facing rotor members radially vaned to a narrow radial parting zone between them to form a fluid working chamber, a driving drum adapted to drive said driving shaft located external to said members and enclosing them, the intervening space within said drum forming a fluid pressure reservoir formed to oppose separation of said members by pressure Within said chamber, a fluid supply system for said chamber including a splined sleeve and a pump driven by said drum and a fluid sump, splined hubs for attaching said members respectively to said shafts, fluid feed passages connecting said pump with said intervening space including feed passages in 12 the splining for the said sleeve connected to the output of said pump, fluid exhaust passages connecting the said working chamber with the said sump and including a, passage in the said driven hub and a central passage in a portion of said driven shaft, and a valve rotating with the said driven hub and located between the said last named hub and shaft passages, said valve being arranged to respond solely to the pressure of said working chamber and that provided by said pump for limiting the pressure to which it responds, in providing an intermittent but continuing flow of fluid between said chamber and said sump during rotation of said driving shaft.

8. In fluid drive devices, the combination of facing vaned rotors, one driving and the other driven, there being included between them a fluid working space opening radially and normally operative to exert pressure axially in a direction to separate the rotors, an overhanging pressuresealed drum, adapted to house said rotors and constituting the pressure feed space open to said working space and arranged to provide fluid pressure opposing the said separating pressure within said working space, and fluid pressure supply and control means effective to provide a body of fluid constantly moving to the interior of said drum to said working space and exhausting therefrom curing the drive interval of said rotors.

- EARL A. THOMPSON.

REFERENCES CITED The following references are of record in the file of this patent:'

UNITED STATES PATENTS Number Name Date 1,270,323 Radclifle June 25, 1918 1,327,080 Brown Jan, 6, 1920 1,640,608 Klimek Aug. 30, 1927 1,766,520 Klimek June 24, 1930 1,873,688 Walker Aug. 23, 1932 2,074,170 Dunn Mar. 16, 1937 2,077,580 Patterson Apr. 20, 1937 2,140,324 Lysholm Dec. 13, 1938 2,144,256 Dufiicld Jan. 17, 1939 2,149,117 Dodge et a1 Feb. 28, 1939 2,154,882 Yoxall Apr. 18, 1939 2,168,862 D Lavaud Aug. 8, 1939 2,168,863 De Lavaud Aug. 8, 1939 2,176,138 Kelley Oct. 17, 1939 2,182,621 Dodge Dec. 5, 1939 2,190,830 Dodge Feb. 20, 1940 2,190,831 Dodge Feb. 20, 1940 2,200,596 Dodge May 14, 1940 2,203,177 Patterson June 4, 1940 2,211,233 Kelley Aug. 13, 1940 2,235,672 Dodge Mar. 18, 1941 2,270,536 Lenning Jan. 20, 1942 

